High frequency vibration test fixture with hydraulic servo valve and piston actuator

ABSTRACT

A vibration test fixture includes a reciprocating slip plate carrying an article subjected to vibration and shock testing. A pair of opposed, spaced apart low oil pressure hydrostatic linear bearings support the slip plate during reciprocating single-axis linear travel on the bearings. The slip plate is mounted to each bearing by a single-axis bearing guide system confining the slip plate to reciprocating longitudinal travel on generally planar, two-dimensional bearing surfaces that support the load of the slip plate. The slip plate is reciprocated by a voice coil-driven hydraulic servo valve and double-acting piston actuator integrated into the vibration table between the bearings that support the slip plate. The voice coil connects directly to a pilot valve in the servo valve assembly. The voice coil generates a vibrating linear motion input that reciprocates the pilot valve at controlled frequencies to induce alternating hydraulic fluid flow control outputs from the servo valve for reciprocating the actuator pistons that drive the slip plate. Other embodiments include a multiple-axis, multiple degree-of-freedom vibration test fixture, and a multiple stage hydraulic servo valve that can be integrated into a piston drive system to achieve high frequency and high force levels in a hydraulic vibration test fixture. Combined large acceleration forces up to about 6 g&#39;s at high frequencies up to about 2000 Hz can be produced in one embodiment of the invention.

CROSS-REFERENCE TO RELATED APPLICATION

This application is a continuation of Ser. No. 08/248,372, filed May 24,1994, now abandoned, which is a division of Ser. No. 07/871,678, filedApr. 20, 1992, now U.S. Pat. No. 5,343,752.

FIELD OF THE INVENTION

This invention relates to vibration testing equipment for simulatingvibration and shock loads on an article under test, and moreparticularly, to a vibration table driven by a hydraulic servovalve-controlled piston actuator for vibrating the table at highfrequencies and/or high force levels. Embodiments of the inventioninclude a double-acting piston actuator with an integrated servo valvecontroller for a hydraulic vibration test table, and multiple-axis,multiple degree-of-freedom vibration test fixtures.

BACKGROUND OF THE INVENTION

The vibration test industry, including all major aerospace firms,automotive manufacturers, electronics companies, and the like, hasadopted use of various methods and systems to simulate vibration andshock environments for determining their products' effectiveness andlongevity when subjected to these environmental extremes. Vibrationtesting is often used to develop vibration-tolerant designs; some use itto verify that a product will survive in its intended vibratingenvironment; others use it to screen out defective parts at an earlystage in the manufacturing process. For instance, vibration and shocktesting often involves the testing of electronic circuit boards adaptedfor use in spacecraft, airplanes, automobiles, etc. The boards aresubjected to vibration testing, often at high frequencies, or high forcelevels, to determine whether or not they will survive the shakingrequired when the hardware is placed in use.

In the past, test fixtures in the form of "vibration tables" have beenused for producing vibration and shock loads on an article under test.These vibration tables (which are also referred to as oil film tables,slip tables, or bearing tables) include a horizontal or vertical tableon which the test article is mounted. The table is vibrated at a desiredfrequency, force level, and/or amplitude during testing. One suchvibration test fixture is manufactured and sold under the name T-FilmTable by Team Corporation, South El Monte, Calif. the assignee of thisapplication. This vibration test fixture is disclosed in U.S. Pat. No.4,996,881, incorporated herein by this reference.

Vibration testing has traditionally been done with the test articlerestrained to move in a single axis. Recent studies show, however, thatvibration testing in three mutually exclusive axes simultaneously cansimulate real world conditions better than single axis testing. Thepresent invention, in one embodiment, comprises a vibration test fixtureadapted for single axis vibration; other embodiments of the inventioncomprise multiple-axis, multiple-degree-of-freedom vibration testfixtures.

Vibration frequency is selected to maximize the effectiveness of thetesting. Transportation tests, for example, require frequencies fromabout 2 to about 500 Hz to simulate truck, rail or air transportationvibration. Screening and qualification tests may require that thefrequency of vibration extend up to about 2,000 Hz for testingcomponents with high natural frequencies. The vibration test generationand control input may be specified to be either sinusoidal motion,random motion, or to duplicate measured real-time waveforms. Vibrationtesting also can be applied at different force levels to meet certainmaximum acceleration specifications.

The different types of test equipment currently available to producevibration motion for a test article have limitations that requireimprovements. Electrodynamic shakers make up the largest percentage ofpurchased vibration force generators. They produce frequency response upto approximately 2,000 Hz and force ranges from about one pound toapproximately 50,000 pounds. However, they are extremely expensive; theyhave relatively small specimen mounting areas; and they must often use a"head expander," a fixture that increases the mounting area for testingin the vertical axis, or a slip table or hydrostatic bearing system fortesting in the horizontal axis. The electrodynamic shaker system addsconsiderable size to the usable specimen mounting surface, often takingup more room on the laboratory or production floor than is needed forthe test object alone.

Hydraulic shakers have proved to be a viable alternative (compared withelectrodynamic shakers) for all vibration testing, except those teststhat require high frequency responses above about 500 to 1,000 Hz.Hydraulic shakers are physically much smaller than electrodynamicshakers, since the conversion of hydraulic power to vibrating motion canbe accomplished with a much smaller mechanism. Hydraulic shakers alsoare much less expensive.

Head expanders, being physical devices, respond to certain drivingfrequencies by resonating. At each resonant frequency, the head expanderdeforms into a characteristic shape, and the frequency and shapetogether define that mode of vibration. The effect of the modes is tomake the vibrating motion on the mounting surface of the head expandernonuniform. At a modal frequency, one location on the mounting surfacemay move less than the shaker input, while a different location may bemoved more than is desired. Head expanders also add mass to thevibration test system, requiring much more force of the shaker than thetest article alone.

A slip plate, and more precisely, the moving element of any horizontalvibration test fixture, has modes of vibration and suffers from the samedeficiencies as head expanders. In addition to normal modes, the slipplate and shaker system is often long enough so that pressure or stresswave phenomena are observed to significantly degrade the uniformity ofvibration of the slip plate surface. The observed phenomenon is that,when controlling the amplitude of a vibration test by monitoring the endof the slip plate, the front and center of the slip plate are often seento have much lower amplitude over a wider range of frequencies.Conversely, if the test is controlled by monitoring the front of theslip plate, then the end is observed to have much higher amplitudes thandesired.

These physical phenomena reduce the useful area of a head expander or aslip plate because it is not possible to obtain uniform vibration inputover the entire surface.

The present invention provides a hydraulic vibration test fixture thatovercomes the disadvantages of prior art electrodynamic shakers andhydraulic shakers. The test fixture avoids the cost and the spacerequirements and limitations of electrodynamic actuators, whileachieving the higher frequency and load level capabilities normallyassociated only with electrodynamic shakers and not achieved by priorart hydraulic vibration test fixtures. Greatly improved uniformity ofvibration amplitude across the useful surface area of the vibrationtable also is produced. Improvements in multiple-axis, multipledegree-of-freedom shakers are also provided.

SUMMARY OF THE INVENTION

Briefly, one embodiment of the invention comprises a vibration testfixture having a fixture base for carrying a unit under test, and ahydraulic vibration actuator having a cylinder, a piston thatreciprocates in a bore within the cylinder, and a reciprocating pistonrod extending from the piston to the outside of the cylinder. Thevibration actuator is mounted to the fixture base to transferreciprocating motion of the piston rod into vibrational motion of thefixture base for vibration testing of the unit under test. The pistonhas an end face, opposite from the piston rod, exposed to a trappedvolume of hydraulic fluid contained within a space in the bore adjacentthe end face of the piston. An inlet port in the cylinder leads to thetrapped hydraulic fluid contained adjacent the end face of the piston. Ahydraulic servo valve adjacent the cylinder has an alternating waveforminput that induces the servo valve to produce a hydraulic fluid flowcontrol output at a frequency corresponding to the desired frequency ofvibration of the unit under test. The hydraulic fluid flow output fromthe servo valve is in close proximity to and passes directly into theinlet port leading to the piston cylinder. Fluid flow from the servovalve output to the cylinder inlet port is over as short a distance aspossible, preferably via a passage of essentially negligible length.This minimizes the volume of compressible fluid trapped between theservo valve output and the piston. This minimized volume of hydraulicfluid between the servo valve and the piston improves hydraulicstiffness, thereby increasing the natural frequency of the system. Onefactor that has limited the frequency response of a hydraulic shaker isthe compressibility of the hydraulic fluid. Flow rate required toachieve frequencies of vibration greatly in excess of the naturalfrequency increases proportional to frequency. That is, a certain amountof additional hydraulic fluid flow within the drive system is requiredto compress the fluid before vibration motion is initiated.

A further embodiment of the invention includes a vibration actuatorhaving a double-acting piston with separate minimized trapped volumes ofhydraulic fluid adjacent corresponding end faces of the pistons. Ahydraulic servo valve can be integrated into the double-acting pistonsso that fluid flow outputs from the servo valve are in close proximityto the trapped volumes of fluid adjacent the pistons. The trappedvolumes of fluid between the servo valve outputs and the input ports tothe piston cylinders are sufficiently minimized to produce highfrequency responses, in one embodiment on the order of 1,000 to 2,000Hz. This high frequency response is possible with good levels of outputforce.

A further embodiment of the invention comprises a servo valve for ahydraulic vibration test fixture. The servo valve has a pilot valvestage in which a pilot valve reciprocates at high frequencies inresponse to a vibrating mechanical frequency input. The servo valve alsoincludes a slave valve stage concentric with and surrounding the pilotvalve stage that amplifies the power output of the servo valve whileminimizing the trapped volume of hydraulic fluid between the two stagesof the servo valve. This results in good power output at highfrequencies. By integrating this servo valve into the piston actuator,the trapped volume of fluid between the pilot valve stage of the servovalve and the end face of the adjacent piston is minimized. In addition,the larger effective area of the slave valve stage spool generates asufficiently high valve flow necessary to achieve high levels of highfrequency response.

Another embodiment of the invention comprises a hydraulic vibration testfixture having a slip plate adapted for carrying an article subjected toshock or vibration testing. Spaced-apart bearings support the slip platefor reciprocating sliding motion along a common longitudinal axis oftravel. Each bearing includes a corresponding bearing guide memberaffixed to the slip plate for guiding the reciprocating sliding travelof the slip plate. A drive actuator for inducing vibrating motion to theslip plate is integrated into the test fixture bearing system. Theactuator is preferably the type of piston actuator described previously,and includes a housing disposed between the spaced-apart bearings, withopposed reciprocating drive members (piston rods) extending fromopposite sides of the actuator housing. The actuator drive members areaffixed to corresponding guide members on the adjacent bearings. Thedrive actuator further includes a hydraulic servo valve controllerintegrated into the actuator housing between the drive members.Vibration motion produced in the drive members by the servo valve ispreferably from the type of concentric pilot-stage/slave-stage servovalve described previously. This induced vibrating motion of the drivemembers causes corresponding reciprocating forces on the bearing guidemembers to vibrate the slip plate. The slip plate is therefore drivendirectly by drive forces originating from within the hydraulic servovalve-controlled actuator that is integrated into the bearing supportand guide system of the test fixture table. This arrangement not onlyproduces high frequency responses in a hydraulic vibration test fixture,but also achieves a greater uniformity of vibration forces spread outover the active surface area of the slip plate.

In one embodiment of the actuator drive, the hydraulic servo valve inputcan be from a voice coil or other transducer directly coupled to thepilot valve for reciprocating the pilot valve at controlled frequencies.The alternating hydraulic fluid flow control outputs from the servovalve are produced in response to the reciprocating motion of the pilotvalve induced by the voice coil. The actuator drive members comprisehydraulically driven pistons extending from opposite sides of theactuator housing, and the servo valve is integrated into the housingbetween the pistons. Since the voice coil, servo valve and pistonactuators are all integrated into a single modular drive unit, thehydraulic fluid flow outputs from the voice coil-controlled servo valvecan be directed toward adjacent end faces of the pistons in a systemthat minimizes the amount of trapped hydraulic fluid between the pilotvalve and the pistons. This arrangement minimizes space requirements andmakes the invention especially applicable to modular actuator-driven,multiple-axis, multiple degree-of-freedom vibration test fixtures.

Another embodiment of the invention comprises a multiple-axis vibrationtest fixture comprising a fixture base for carrying a unit under test,and a plurality of separate vibration actuators affixed to the base.Each of the vibration actuators preferably comprises the modularactuator-servo valve unit described previously. Each vibration actuatorincludes an actuating arm that reciprocates along an exclusive axisaligned at an angle with respect to and intersecting the axes of othersimilar vibration actuators. The vibration actuators apply a vibratingmotion to the fixture base along their corresponding axes to vibrate thefixture base in multiple axes. Means are provided for decoupling fromeach vibration actuator the motion imparted to the fixture base from theother vibration actuators. The forces from other vibration actuators aredecoupled from each actuator preferably by spherical bearing ends of thepiston rods which are slidable and rotatable with respect tocorresponding spherical bearing supports affixed to the fixture base. Byproviding such freedom of sliding and rotating motion with respect tothe force-applying ends of each vibration actuator, the modularvibration actuator units can be mounted to a fixture base to applyvibration forces in multiple directions in many different configurationsof multiple axis, multiple degree-of-freedom vibration test fixtures.

The vibration test fixture provided by the integrated actuator drivesystem provides substantial improvements over the conventional vibrationtesting apparatus described previously. These improvements include lessspace required, easier access around the vibration test table, higherfrequency response, reasonably uniform vibration levels over the entireuseful surface area of the vibration table, reduced cost, and greatlyreduced oil leakage problems. The objective of obtaining high frequencyresponses in the range from about 1,000 Hz to about 2,000 Hz in ahydraulic shaker is achieved by the system minimizing the trapped volumeof compressible fluid between the piston and valve, in combination withproducing the high valve flow rate necessary to operate withcompressible flow at such high frequencies.

In addition, the invention enhances modular design of vibration tablesand their drive systems. The use of multiple actuator drive systemmodules, for instance, can produce a corresponding increase in actuatorforce when large single axis g-forces are required. The modular designof the actuator drive system also provides test fixtures for vibrationtesting in mutually exclusive axes, in addition to single-axis vibrationtesting.

These and other aspects of the invention will be more fully understoodby referring to the following detailed description and the accompanyingdrawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is an exploded semi-schematic perspective view illustratingcomponents of a vibration test fixture according to principles of thisinvention.

FIG. 2 is a fragmentary plan view showing a hydraulic servo valve anddouble-acting piston actuator integrated into the test fixture between apair of spaced-apart bearings that support a vibration table.

FIG. 3 is a side elevation view taken on line 3--3 of FIG. 2.

FIG. 4 is an enlarged top plan view of the apparatus shown within thecircle 4 of FIG. 2.

FIG. 5 is a cross-sectional view illustrating components of the servovalve.

FIG. 6 is a semi-schematic view showing a voice coil and pilot valvecomponent of the servo valve.

FIG. 7 is a semi-schematic view showing a slave valve spool component ofthe servo valve.

FIG. 8 is a schematic view of a prior art vibration actuator and servovalve system described in the Example set forth in the detaileddescription.

FIG. 9 is a semi-schematic front elevation view showing one embodimentof a multiple-axis, multiple degree-of-freedom vibration test fixture.

FIG. 10 is a semi-schematic top view taken on line 10--10 of FIG. 9.

FIG. 11 is a semi-schematic front elevation view showing an alternativeembodiment of a multiple-axis, multiple degree-of-freedom vibration testmodule.

FIG. 12 is a semi-schematic side elevation taken on line 12--12 of FIG.11.

FIG. 13 is a semi-schematic top view taken on line 13--13 of FIG. 11.

FIG. 14 is a semi-schematic perspective view of a multiple-axis drivesystem of the test module shown in FIGS. 11-13.

DETAILED DESCRIPTION

FIG. 1 is an exploded perspective view showing principal components of avibration test fixture which illustrates one embodiment of theinventions. FIGS. 2 through 7 are more detailed views of a specificembodiment of the invention shown generally in FIG. 1. The fixture shownin FIGS. 1 through 7 is one embodiment of the invention incorporatedinto a hydraulic shaker table vibrated along a single horizontal axisfor vibration and shock testing of an article affixed to the table. Thisembodiment is for the purpose of illustrating one use of the invention;other uses also are possible, such as in vertical axis shakers ormultiple-axis shaker tables. FIG. 1 also illustrates a single hydraulicdrive actuator unit for applying vibration forces along a single axis ofthe table, but this is simply one embodiment to illustrate principles ofthe invention. Multiple drive systems of similar design can be used onmultiple parallel drive axes or on multiple exclusive axes for vibratingthe table.

In addition, the invention is illustrated in FIGS. 1 through 7 withrespect to a single axis horizontal shaker table, preferably of the typemanufactured by Team Corporation, and sold under the name T-Film Table.As mentioned previously, this type of vibration test fixture isdescribed in U.S. Pat. No. 4,996,881. The invention can, however, beused in other types of vibration tables.

To briefly describe the type of shaker table with which the invention isused, as illustrated in FIG. 1, the test fixture includes a rigid andgenerally planar slip plate 10 mounted for reciprocating linear travelon flat upper bearing surfaces 12 of a pair of longitudinally spacedapart support bearings 14. The slip plate 10 is of generally uniformthickness and has a generally rectangular surface area within a regionof its bottom surface that is supported for linear travel on the upperbearing surfaces 12. Recessed regions (shown at 51 in FIG. 2) arelocated in the top surfaces of the bearings 14. These recessed regionsare used to supply lubricating oil under low pressure to the bottomsurface of the slip plate and to the tops of the bearings forlubricating sliding travel of the slip plate 10 back and forth on theupper bearing surfaces 12 during vibration testing.

According to principles of the invention, the vibration table is drivenby a hydraulic actuator drive unit 16 having a double-acting hydraulicdrive piston actuator controlled by a hydraulic servo valve 18. Theactuator drive unit 16 is disposed between the pair of axially alignedand spaced-apart guide bearings 14 that support the shaker table 10 forits linear reciprocating travel. The actuator drive unit includespistons 20 and 22 for applying a reciprocating linear force to the slipplate for vibrating the slip plate during use. Although the invention isillustrated with respect to an actuator having a pair of pistons forproducing output forces in opposite directions, this is one embodimentof the invention; the invention also can be adapted for use with asingle piston actuator.

The drawings (see FIGS. 2 and 3) illustrate one embodiment of thevibration test fixture, in which the bearings 14 are rigidly affixed tothe upper surface of a steel base plate 24, preferably of rectangularconfiguration. An upright outer wall 26 extends around the rectangularouter perimeter of the base plate to provide a means for retaininglubricating oil within the fixture. A manifold (not shown) mounted toone edge of the base plate provides for filtration and recirculation oflubricating oil through the bearings.

The bearings 14 at opposite ends of the drive actuator unit preferablycomprise separate three-layer structures, which are shown in more detailin U.S. Pat. No. 4,996,881. Each bearing has a generally rectangularbottom plate; a pair of laterally spaced-apart, long and narrow,generally rectangular middle plates extending parallel to each otheralong opposite sides of the bottom plate; and a pair of laterally spacedapart, generally rectangular top plates extending parallel to each otheralong opposite sides of the middle plates. The narrow middle plates forma wide, shallow bottom portion of an inverted T-shaped channel 28extending along the central axis of each bearing. The top plates arewider than the middle plates, and are spaced apart by a narrow gap 30which forms the narrow portion of the inverted T-shaped channel in eachbearing. The inverted T-shaped channels are aligned along a common driveaxis of the fixture. The flat upper bearing surfaces 12 are formed atthe tops of the top plates and lie in a common flat plane to providebearing support for the bottom of the slip plate 10. A rectangular topsurface 32 of the actuator drive unit 16 is aligned in a common flatplane with the upper surfaces 12 of the bearings 14 to provide acontinuous flat supporting surface for the slip plate.

The slip plate is mounted for guided single axis travel relative to thebearings by separate inverted T-shaped guide members 34 rigidly affixedto the bottom of the slip plate. These inverted T-shaped guide membersare slidably engaged in corresponding inverted T-shaped channels 28 inthe bearings 14 for guiding the slip plate 10 during use. Each guidemember 34 is a two-component structure which includes a relatively widerlower guide member bearing block 36 that slides in the wide bottomportion of the T-shaped channel, and a relatively narrower upper guidemember bearing block 38 of rectangular configuration which slides in thenarrow upper portion 30 of the T-shaped channel. Each upper bearingblock is affixed to a corresponding lower bearing block and is alsorigidly affixed to the bottom of the slip plate. Threaded holes 40 ineach bearing guide member 34 and cooperating threaded holes 42 in theslip plate receive fasteners for rigidly affixing the bearing guidemembers to the slip plate. The bearing guides 34 provide means forguiding single-axis travel of the slip plate along the inverted T-shapedchannels 28 in the bearings.

Oil flow ports 44 open through the working faces of each bearing blockand the T-shaped channel in each bearing block. This provides a meansfor supplying a film of lubricating oil to the lateral and verticalrestraint surfaces of the bearing for lubricating travel of the T-shapedguide member 34 relative to the channel 28 in each bearing, togetherwith means for supplying a film of lubricating oil to the upper bearingsurface 12 for lubricating reciprocating longitudinal travel of the slipplate 10 on the bearing surfaces.

The hydraulic servo valve 18 and double-acting pistons 20 and 22 arecontained in a modular unit integrated into the slip table bearingsystem, by mounting the actuator module 16 in-line between the bearings14 that guide and support the slip plate 10. The pistons 20 and 22contained in the hydraulic valve actuator reciprocate on a common axis45 (see FIG. 1) and extend outwardly from opposite sides of the actuatorhousing 46. The pistons are aligned and engaged with the T-shapedbearing guide members 34 so that reciprocating travel induced in theactuator pistons causes alternating thrust forces from the pistons todrive the bearing guide members in a corresponding reciprocating motionalong the bearing guide axis of travel. The hydraulic servo valve 18controls piston movement by providing a controllable frequency input forinducing vibration motion directly to the pistons. The servo valve unit18 contains a transducer such as a voice coil unit 47 connected directlyto a pilot valve 48 (shown best in FIGS. 5 and 6) for driving the pilotvalve at controllable frequencies set by an electronic voice coil driveinput signal 49 (see FIG. 6). The pilot valve motion produces hydraulicfluid flow control outputs C₁ and C₂ (see FIG. 5) at the controlledfrequency from the servo valve. The fluid control outputs from the servovalve are directed toward the faces of the pistons 20 and 22 to vibratethe pistons at the frequency set by the input signal 49. The hydraulicservo valve 18, the double-acting piston actuator, and the voice coildrive unit 47 are all integrated into a common module 16 that isdirectly integrated into the slip plate bearing guide system andpositioned beneath the center of the slip table. This avoids the priorart use of external drive means, such as electrodynamic actuators, or anexternal single axis hydrostatic bearing actuator drive common in priorart hydraulic slip tables. Other improvements are also provided asdescribed in more detail below.

FIGS. 2 and 3 show the integrated servo valve and double-acting pistonactuator module 16 mounted to the base plate 24 of the test fixturebetween the spaced-apart pair of T-shaped bearing guide members 34aligned on the axis of vibration 45. The outer housing 46 of theactuator is mounted in-line between the T-shaped bearings 14 thatsupport opposite sides of the slip plate. As mentioned previously, theactuator housing 46 has a flat rectangular top surface 32 at the sameelevation as the flat rectangular top surfaces 12 of the bearings 14, toprovide a continuous, flat top surface for supporting the load of thereciprocating slip plate 10. The actuator housing 46 includes ports 50for supplying low pressure lubricating oil to the top surface of thehousing to provide a lubricating film for the undersurface portion ofthe slip plate that rests on top of the actuator housing.

The pistons 20 and 22 are aligned along the common axis of vibration 45,and opposite end portions of the pistons project outwardly from oppositesides of the actuator housing 46. The pistons each reciprocate incorresponding cylinders 52 contained internally within the actuatorhousing. The outer end portions of the pistons have spherically curvedthrust-applying bearings 54 that apply pressure to matching recessedspherically curved bearing surfaces of piston guide cups 56 affixed tocorresponding T-shaped bearing guide members 34 aligned with the axes ofthe pistons. The hydraulic fluid pressure within the system causes thespherical ends of the piston bearings to apply pressure to the sphericalbearing cups 56. The spherical ends of the pistons are otherwise free toslide or rotate relative to the spherical bearing cups, which provides aself-alignment function for the pistons. The pistons are shown asuniform diameter members with enlarged spherical bearings at the ends;alternatively, the pistons can have reduced diameter piston rodsprojecting from opposite ends of the actuator housing with separateenlarged spherical bearings 54 at the ends of the piston rods.

FIGS. 2 through 4 best illustrate the hydraulic servo valve assembly 18incorporated into the piston actuator housing 46. The servo valveassembly, described in more detail below, includes the pilot valve spool48 (see also FIG. 6) which is contained within the servo valve unit 16and mounted within the actuator housing 46 along an axis 58perpendicular to the piston axis of travel 45. The pilot valve spool 48is driven at controllable frequencies by a voice coil 60 containedwithin the voice coil unit 47. An electromagnetic field 62 for the voicecoil is also contained within the voice coil unit 47 adjacent the voicecoil 60. The electronically controlled drive input signal 49 (see FIG.6) to electromagnetic field coil 62 induces mechanical vibrationalmotion in the voice coil. The pilot valve spool 48 of the servo valveassembly 18 is directly connected to the voice coil output, andoscillating motion induced in the voice coil is transmitted directly tothe pilot valve spool for reciprocating the pilot valve spool at thesame frequency. Vibration of the pilot valve is along the axis 58perpendicular to the piston axis of vibration. The input drive signal 49can be a sinusoidal input, a random motion input signal, or it canduplicate measured real-time waveforms. The voice coil represents anexample of a transducer which oscillates in response to an alternatingcurrent in a magnetic field to generate a vibration motion of the pilotvalve along its axis; other similar transducers could be used for thesame purpose.

As the pilot valve reciprocates along the axis 58 perpendicular to thepiston axis of travel 45, the servo valve unit 18 produces resultinghydraulic fluid flow control outputs C₁ and C₂ which flow directly intolow volume hydraulic fluid pressure chambers 64. The chambers 64 extendfrom the control outputs of the servo valve to the end faces of thepistons 20 and 22 in the cylinders 52. The resultant forces of thehydraulic fluid flow control outputs C₁ and C₂ (see FIG. 5) are directedoutwardly from the servo valve output ports along axes parallel to thedrive axis of the slip table (and perpendicular to the end faces of thepistons); the alternating fluid flow outputs from the servo valve applythese thrust forces to the ends of the pistons at the same frequency asthe frequency induced in the pilot valve spool 48 by the voice coil 60.In one embodiment, voice coil input drive signals above about 1,000 Hzand as high as about 2,000 Hz generate corresponding hydraulic fluidflow control outputs directed at the pistons at the same frequency tovibrate the test fixture slip plate at the same frequency level. The lowvolume hydraulic fluid chambers 64 are shown exaggerated in size in FIG.4. The passages 64a from the servo valve outlets (at C₁ and C₂) to thecylinder inlet ports 64b are also exaggerated in size for clarity. Thehigh frequency response of the piston actuator is achieved primarilybecause the system is designed to minimize as much as possible thevolume of trapped compressible hydraulic fluid between the outlet portsC₁ and C₂ of the servo valve and the end faces 65 of the pistons 20 and22 which are exposed to the volume of hydraulic fluid contained in thebore of each cylinder. To minimize the volume of trapped fluid, theforce-applying ends of the pistons extend outwardly in oppositedirections from the piston actuator housing, and the end faces 65 of thepistons, which are opposite from the force-applying ends of the pistons,are positioned adjacent to one another and are exposed to the separatecontrol chambers 64 of hydraulic fluid which are in direct internalcommunication with the outlet ports of the servo valve. In the preferredembodiment, the servo valve is positioned directly between and in closeproximity to the control chambers 64 so that fluid flow control outputsfrom the servo valve are passed directly into the low volume controlchambers 64, while minimizing the total volume of compressible hydraulicfluid between the valve and piston. The invention avoids long and narrowpassages or conduits for hydraulic fluid flow between the servo valveand the bore of the piston cylinders. Hydraulic fluid is a compressiblesubstance, and its compressibility in such long, narrow passageways cangreatly reduce the high frequency response of the actuator. Byminimizing the trapped volume of fluid, which is the main compliantelement in the system, the actuator achieves high frequency response.

More specifically, the servo valve outlets are in close proximity to theinlet ports of the cylinders, and the trapped volume of fluid betweenpiston end faces 64 and servo valve outlets C₁ or C₂ is minimized asmuch as possible. The fluid flow from the valve to the piston istherefore as direct as possible; the length of each passage 64a is asshort as possible, and is preferably of essentially negligible length.The volume of fluid within the cylinder adjacent the piston also is aslow as possible, while maintaining sufficient fluid pressure onsubstantially the entire end face area of the piston. In one embodimentin which the pistons have a diameter of about two inches, the volume offluid within each cylinder (adjacent the piston end face 65) occupiesfrom nearly zero to about two inches of stroke length, or an average ofabout one inch. Actual stroke length of the piston operating at highfrequencies over about 1000 Hz is about 0.1 inch. The length of thepassage 64a is less than about one inch, i.e., less than the radius ofthe piston, more preferably a fraction of an inch, and as a practicalmatter is as short as possible.

FIG. 5 is a cross section of the hydraulic servo valve assembly 16 whichincludes the voice coil 60 rigidly affixed to one end of the pilot valvespool 48 which extends along the center line of the servo valve housing.As mentioned previously, the pilot valve spool 48 reciprocates linearlyalong the longitudinal axis 58 of the servo valve assembly directly inresponse to corresponding vibration of the voice coil. The pilot valveports to a hydraulic pressure supply contained in separate hydraulicaccumulators 66 contained in the actuator housing on opposite sides ofthe servo valve assembly. The hydraulic fluid for producing the servovalve outputs circulates in a closed hydraulic fluid system that portsin and out of the servo valve to and from the hydraulic accumulators.The servo-valve is a four-way valve assembly in which the pilot valvespool is concentrically mounted for reciprocating sliding movement in anelongated cylindrical bore 68 inside a stationary pilot valve sleeve 70.Coil springs 72 apply a spring biasing forces to the ends of the movingpilot valve spool. Hydraulic fluid at pressure P_(in) passes throughinput ports in the pilot valve spool and along a small diameterconcentric bore in the pilot valve spool. The fluid at a pressureP_(out) flows out from pressure output ports near the opposite end ofthe bore through the pilot valve spool. Sections 78 of reduced diameteradjacent opposite outer ends of the flow edges provide the inlet andoutlet openings of the inlet and outlet ports. A reduced diametercentral portion 80 of the pilot valve spool between the flow edges 78opens into a pilot valve hydraulic fluid supply return chamber. The flowedges of the pilot valve spool communicate with cylindrical recessedareas 82 inside the pilot valve sleeve for controlling the cycling anddirection of hydraulic fluid pressure control outputs C₁ and C₂ of theservo valve.

The servo valve assembly also includes a slave stage that provides poweramplification for the servo valve. The slave stage includes a movableslave valve spool 84 (also shown in FIG. 7) concentric with the servovalve axis 58 and disposed concentrically around the outside of thepilot valve sleeve 70. The slave stage also includes a stationary slavevalve sleeve 86 concentric with and extending around the exterior of theslave valve spool. The slave valve spool has an internal axial bore 88that slides on a cylindrical outer surface 90 of the stationary pilotvalve sleeve 70. The slave valve spool is adapted for spring biasedlongitudinal reciprocating motion inside the servo valve in response tothe alternating hydraulic fluid control pressure outputs induced withinthe servo valve by the vibrating motion of the pilot valve spool. Suchinduced motion of the slave valve spool produces the servo valve controlpressure outputs C₁ and C₂ as described in more detail below.

The slave valve spool 84 has longitudinally spaced-apart enlargedcylindrical sections which include first and second end bearings 92 thatsupport sliding travel of the slave valve spool inside a central bore 94within the stationary slave valve sleeve 86. The ends of the slave valvespool are biased by corresponding coil springs (not shown). First andsecond cylindrical flow edges 96 are located near the central portion ofthe slave valve spool. A region 98 of reduced diameter between the flowedges of the slave valve spool opens into a return chamber 100 thatcommunicates with a hydraulic fluid return 102 from the pilot valve. Thehydraulic fluid return from the pilot valve to the return section of theslave valve sleeve passes from lateral return ports 104 in the pilotvalve sleeve, through communicating lateral return passages 106extending through the slave valve spool, and through communicatinglateral return ports 108 in the slave valve sleeve. These return portscommunicate with return ports that lead to the supply section of theslave valve sleeve.

Hydraulic fluid under pressure P₁ and P₂ is supplied from the returnsystem to the slave valve section through annular passages 110 in theouter housing surrounding the slave valve sleeve 86. Hydraulic fluidunder pressure P₁ and P₂ is supplied to the moving slave valve spool 84through lateral supply passages 112 extending through the slave valvesleeve to recessed annular supply pressure regions 114 of the slavevalve spool adjacent the first and second flow edges 96 of the slavevalve spool. Hydraulic fluid under pressure communicates between theflow edges of the slave valve spool and the pressure input ports to thepilot valve through lateral passages 116 extending from the recessedannular region of the slave valve spool, through the body of the slavevalve spool and the pilot valve sleeve, and to the pressure input P_(in)of the pilot valve 48.

Alternating hydraulic fluid pressure control outputs C₁ and C₂ from theservo valve slave stage are produced along spaced-apart recessed annularpassages 118 in the outer housing that communicate with correspondingrecessed annular control passages 120 in the stationary slave valvesleeve. The hydraulic fluid pressure control outputs C₁ and C₂ aredirected to output stages of the servo valve for direct communication ofhydraulic fluid flow from the servo valve to the pistons that are drivenin response to fluid flow cycling of the servo valve. Hydraulic fluidunder pressure in the recessed pressure regions 120 of the slave valvesleeve communicate with first and second lateral passages in the slavevalve sleeve 122 leading to the control output regions C₁ and C₂ of theservo valve. The alternating flow control outputs C₁ and C₂ arecontrolled by the flow edges 96 of the slave valve spool 84 whichreciprocates to open and close the supply of hydraulic fluid underpressure to the outputs C₁ and C₂ of the servo control valve. In theclosed hydraulic fluid pressure system, axial cycling of the slave valvespool in response to corresponding cycling of the pilot valve spoolalternately opens and closes the control passages 120 to the controloutput regions of the servo valve. A pressure output from control regionat C₁ simultaneously produces a reverse flow of control pressure fromthe control region at C₂ to the return path, and vice versa.

During reciprocating motion of the pilot valve, hydraulic pressure actsalternately on differential areas shown at P_(A) and P_(B). Hydraulicfluid under pressure reaches pressure area P_(A) through lateral outletpassages 124 in the pilot valve sleeve and 126 in the slave valve spool.Hydraulic fluid under pressure reaches pressure area P_(B) throughlateral passages 128 in the pilot valve sleeve and 130 in the slavevalve spool. The differential pressure at areas P_(A) and P_(B) causescycling of the slave valve at the same frequency as the pilot valve.Axial reciprocating travel of the slave valve alternately opens andcloses the lateral control passages 120 in the slave valve sleeve toreceive or to block off hydraulic fluid pressure in the lateral pressurepassages. This causes hydraulic fluid flow at control outputs C₁ and C₂to cycle at the frequency input from the voice coil to the pilot valve;and these control outputs are directed to the end faces of the pistonsin a low volume hydraulic fluid pressure circuit that cycles the pistonsat the frequency produced by control outputs C₁ and C₂.

Thus, vibrating longitudinal single-axis motion induced in the pilotvalve by the voice coil causes hydraulic fluid under pressure to flow ascontrol outputs in a closed hydraulic fluid system in which the controloutputs are alternately directed to the pistons 20 and 22 in small,short, low volume pulses at the same frequency as the motion induced bythe voice coil. In addition, the slave valve stage of the servo valve isconcentric with the pilot valve stage. This arrangement avoids longnarrow, passages or lines between separate pilot valve and servo valvescharacteristic of the prior art. This minimizes the compliancy effectsof hydraulic fluid between the input stage of the servo valve and itsoutput. By minimizing the compliancy of the hydraulic fluid contained inthe servo valve, better high frequency response is achieved. The largesize (area) of the servo valve slave stage spool also increases valveflow rate substantially. This increased valve flow combines with thereduced compliancy (of the trapped compressible fluid) to achievevibration response at high frequencies. The concentric arrangement ofthe pilot stage and the slave stage also provides a much smaller servovalve unit which can be incorporated effectively into the pistonactuator housing so that both outputs from the servo valve are fromopposite sides of the servo valve and directly into the low volume fluidcontrol chambers adjacent the end faces of the pistons.

The hydraulic vibration test fixture has the following advantages. Theactuator pistons drive the moving elements of the bearings both in frontof and behind the actuator instead of at the edge of the slip plate. Thesingle actuator thus inputs its force to the slip plate through the twoslip plate/T-film bearing connections, spreading the force input over alarge area. Multiple small hydraulic shakers can be used tosimultaneously provide high frequency response and input the vibrationforce uniformly to the entire table surface. The shakers are modular,including both the servo valve and piston in a single body. The shakercan be mounted in a common sump and has no oil seals. The oil is allowedto escape from the actuator assembly and can be collected in the sump.The sump can be sealed so that contaminants cannot get in or out. In thecase of the horizontal vibration table system shown in the drawings, inwhich the shaker is used in conjunction with the T-bearings, theactuator is configured to fit the T-film bearing envelope so that it canreplace the bearing in a T-film table assembly directly. The top surfaceof the actuator has an oil film bearing similar to the T-film bearing tolubricate and support the slip plate. The shaker shafts provide lateralmotion restraint and the spherical couplings allow the shakers to go outof phase without damage. The double piston design ensures a preloadedload path, eliminating backlash, and it allows the valve assembly to bemounted between the pistons with extremely short oil passages from thevalve to the pistons. This enhances the high frequency response of theactuator system. The hydraulic accumulators are integral with the shakerbody. All oil porting is internal and are no external oil lines. Thepistons drive the table or T-bearing elements through spherical padbearings. This has the benefits of the actuator driving either a singleaxis table or a multiple degree of freedom table. It also eliminatesalignment as a critical parameter during assembly. The integrated shakeruses the high power-to-size ratio of hydraulic shakers and can producevibration above 1,000 Hz and up to 2,000 Hz at high acceleration levels.In one embodiment the actuator produces 3,500 pounds of force rms for arandom signal input from 20 to 2,000 Hz. Higher force levels can beproduced with multiple actuator modules aligned to apply the force in acommon direction.

EXAMPLE

The dynamic performance of an electro-hydraulic servo system may bequantified with respect to three parameters: hydraulic naturalfrequency, required hydraulic flow rate, and freedom fromnon-linearities. All of these factors are strongly affected by thedistance between the servo valve and the actuator piston. It may beshown that performance is greatly enhanced by making this distance asshort as possible.

To illustrate this point quantitatively, consider the following twosplit actuator systems. The first is a prior art design illustrated inFIG. 8, in which the two half actuators bear on opposite sides of aprojection from the shake table, with the servo valve located at thecenter and connected by pipes or hoses to the far ends of the actuators.The second system, the subject of this invention, has the two actuatorsplaced back to back, with the servo valve connected to the pistons bypassages of negligible length.

Assume that the pistons are 2 inches in diameter, with a maximum strokeof ±0.5 inches. The system has peak velocity of 50 inches/second and apeak pressure of 3,000 psi. The displacement and velocity correspond tothose commonly achieved by conventional electrodynamic shakers. Assumefurther, that the fluid velocity in the lines is limited to 150inches/seconds, thus requiring a cross sectional area of 1/3 squareinches. Finally, assume that the distance between the cylinder end andthe servo valve outlet in the first case is 12 inches, and in the secondcase essentially zero.

The first parameter, the hydraulic stiffness of the actuator is:##EQU1## where: K=spring rate,

A=piston area,

B=fluid bulk modulus,

s=stroke, and

V'=volume in the connecting lines.

For the first actuator, the effectie bulk modulus, including line andcylinder compliance, is found to be about 100,000 psi. Thus, ##EQU2##

For the second case, the effective bulk modulus is about 200,000 psi, soK2=2,500,000, or a factor of more than seven times greater. SinceNatural Frequency is proportional to the square root of the stiffness,the new system will have more than 2.6 times the natural frequency ofthe old system and may thus be presumed to be controllable within ahigher frequency range.

For example, assume that the moving mass is 300 pounds, so that the peakacceleration is about 31 G, a performance achievable by a goodelectrodynamic shaker. The hydraulic natural frequency of the new shakerwill be about 280 Hz, while it is only 107 Hz for the old one.

The second parameter, the required flow rate, is determined as follows.It is common practice to drive electrohydraulic shakers to frequenciesmuch higher than their natural frequency in order to do effectivevibration testing. However, a limit to the high frequency performance isset by the compressibility of the flow, that is, the extra flow requiredto compress the fluid if the piston is held stationary. The compressibleflow becomes the dominant flow at high frequencies (above the naturalfrequency) where the piston motion is essentially zero.

The compressible flow is:

    Q.sub.c =VPF/B

where:

V is the total trapped volume,

P is the peak pressure, and

f is the sinusoidal excitation frequency.

The volumes of the new and old configurations are (π) and (π+8) or 3.14and 11.14 cubic inches, respectively. The compressible flows at 1,000 Hzthen are 47.1 and 334 cubic inches per second (cis) respectively.(Assuming the same effective bulk moduli as before.) The old designrequires seven times the flow of the new design, and therefore a valvecapable of flowing seven times as much oil. An "avalanche effect" comesinto play as the larger servo valve itself has lower frequency responseand the greater flow requires larger passages, which, in turn, add tothe compressible flow requirement. Thus, the possibility of reachinghigh frequencies diminishes with every increment of added volume betweenthe servo valve and the actuator. This is avoided in the new design andis fundamental to its higher frequency response.

The third parameter of performance, the system non-linearity, is relateddirectly to the distance between the valve and the pistons (as opposedto the volume between them). Because of the finite velocity of sound inthe fluid, a pressure pulse originating at the servo valve will travelto the actuator and back to the valve in a finite time. A series of suchpulses making up a sinusoidal excitation will excite a frequency of:

    f=Vs'/2

where:

V is the velocity of sound, about 35,000 to 50,000 inches/second, and

s' is the distance between valve and the piston.

For the old system, with s'=12 inches, this results in an apparentfrequency of about 1,500 Hz. This standing pressure wave shows up asnoise in the acceleration of the actuator (and therefore on the testspecimen). In the new system, with s'=1 inch, the pressure pulsefrequency is about 17,500 Hz. No standing waves or extraneous noise willbe produced in the frequency range normally of interest in mechanicalvibration testing.

FIGS. 9 and 10 schematically illustrate an alternative form of theinvention comprising a multiple-axis, multiple degree-of-freedomvibration test fixture 130. The fixture includes a normally horizontalupper fixture base 132 for carrying a unit under test. The base 132 issupported by three sets of upwardly extending pairs of rigid supportarms 134, 136 and 138 aligned at separate angular orientations aroundthe bottom of the upper base 132. The support arms are mounted to arigid stationary lower base 140 so as to accommodate vibration motion ofthe upper base of the fixture in both the vertical plane and thehorizontal plane. A separate piston actuator module is affixed to eachsupport arm system. Each piston actuator module is aligned on an axisextending at an angle with respect to the axis of the other two modules.Thus, as shown best in FIG. 10, a first module 142 is aligned along anaxis 144, a second module 146 is aligned along an axis 148 and a thirdmodule 150 is aligned along an axis 152. Each module includes a pair ofpiston actuators combined as a unit in a separate housing, with one setof pistons reciprocating along a horizontal axis and a second set ofpistons reciprocating along a vertical axis. FIG. 9 best shows the firstmodule 142 having horizontal piston actuators 154 and 156 and verticalpiston actuators 158 and 160 extending outside a common housing 162along mutually orthogonal axes. FIG. 10 shows the horizontal pistonactuator pairs for each of the three modules, including piston actuators164 and 165 for module 146 and piston actuators 166 and 167 for pistonactuator module 150. Vertical piston actuators are shown in FIG. 10, andthese include the piston 158 of the first module 142, a vertical piston168 for module 146 and a vertical piston 170 for module 150.

Each of the three modules preferably comprises a hydraulic servo valvecontrolled-double acting piston actuator similar to the servovalve-piston actuator shown in FIGS. 1 through 7. Thus, each unitincludes a separate voice coil (shown schematically at 172 in FIG. 10)for providing an oscillating input signal to each servo valve of eachpiston actuator module. There are two voice coil operated hydraulicservo valves for each unit, one for the horizonal pair of pistons andone for the vertical pair of pistons. The hydraulic servo valve andvoice coil controllers for each module are preferably the concentricpilot valve-slave valve arrangement similar to that shown in FIG. 5.

The outer working ends of the four piston rods extending to the outsideof each housing of each module are engaged with rigid portions of thefixture support system. Thus, referring to FIG. 9, the ends of thehorizontal pistons 154 and 156 are engaged with opposite support arms134, the end of the vertical piston 158 is engaged with the bottom ofthe rigid upper support base 132, and the end of the vertical piston 60is engaged with the bottom of the u-shaped structure which includes theupright support arms 134. The ends of the piston rods in the othermodules 146 and 150 are similarly affixed to the fixture support system.In each module the connections between the ends of the pistons and thefixture support system is through a cooperating pair of sphericalbearings similar to the spherical bearing connections illustrated inFIGS. 1 through 4. These spherical bearing connections are illustratedschematically at 174 in FIGS. 9 and 10. In each spherical bearingconnection the hydraulic servo system applies fluid pressure outwardlyon the piston arms to hold them in pressure contact with theircooperating support bearings. Thus, the spherically curved bearing endsof the piston rods are able to slide and rotate relative to theircorresponding spherical bearing connections to the fixture supportsystem. As a result, vibrational forces acting along any of the threeaxes of the vibration modules are decoupled from motion induced on themfrom vibration forces induced on the fixture from vibration of either orboth of the other two vibration modules.

During use of the vibration fixtures shown in FIGS. 9 and 10, the pistonactuator pairs in each module are actuated to oscillate along theirrespective axes in accordance with any desired waveform test signal.This permits the vibration table to be vibrated along orthogonal x andy-axes and rotated with a z-axis rotation in the horizontal plane bycorresponding combinations of vibration of the three sets of horizontalpiston actuators. Separately, components of vertical forces are inducedby the three sets of vertical piston actuators to oscillate the table invarious combinations of pitch, yaw or heave orientations in the verticalplane, combined with horizontal plane movement. This provides amultiple-axis, multiple degree-of-freedom vibration test fixture.

FIGS. 11 through 14 illustrate an alternate form of the inventioncomprising a multiple-axis, multiple-degree-of freedom vibration testfixture 150 which includes a cube shaped rigid outer fixture 152 havingend faces in six mutually orthogonal planes. Separate units under testcan be affixed to the cube fixture base on any of the five exposed endfaces for vibrational testing as desired. The cube fixture is actuatedby a hydraulic servo valve-controlled multiple piston actuator 154contained within the interior of the cube fixture. The multiple pistonactuator is best shown in FIG. 14. This piston actuator includes sixseparate pairs of piston actuators aligned along six separate axes forinducing vibration motion to the cube fixture base to oscillate thefixture with six degrees of freedom. In the illustrated arrangement, thesix pairs of piston actuators are mounted on a common central actuatorsupport 156. The actuators include an upper horizontal piston actuator158 and a lower horizontal piston actuator pair 160 extending above oneanother in a common vertical plane; a pair of horizontally extending andlaterally spaced-apart piston actuator pairs 162 and 164 on oppositesides of the actuator unit lying in a common horizontal plane; and apair of laterally spaced-apart upright piston pairs 166 and 168 onopposite sides of the actuator unit extending in a common verticalplane.

Each piston actuator pair is similar to the double-acting pistonactuator illustrated in FIGS. 1 through 4, and each is controlled by aseparate voice coil-driven concentric servo valve unit similar to thatshown in FIGS. 5 through 7.

Further, the piston actuators of each actuator pair have correspondingspherical bearing connections similar to those described above andillustrated schematically at 170. As mentioned previously, thesespherical bearing connections decouple the vibrational motion of eachactuator pair unit from the motion induced on the fixture in otherdirections by simultaneous operation of the other vibration testfixtures. Thus, the vibration test fixture can be vibrated in multipleaxes with multiple degrees of freedom.

What is claimed is:
 1. A high frequency linear hydraulic servovalve-actuated vibration test assembly for producing alternatinghydraulic fluid flow control outputs the servo valve-actuated testassembly comprising:an energy input source comprising a transducer toreceive an alternating input signal of a selected frequency and convertthe input signal to linear mechanical vibrational motion; a servo valvehaving a pair of fluid flow control output ports and in which an outputflow of hydraulic fluid is produced from the control output ports of theservo valve; the servo valve including a movable hydraulic pilot valvespool that receives the vibrational motion from the transducer to applylinear reciprocating motion to the pilot valve spool along an axis; amovable hydraulic slave valve spool surrounding and concentric with thepilot valve spool; a fixed pilot valve sleeve surrounding and concentricwith the pilot valve spool, the pilot valve sleeve located between thepilot valve spool and the slave valve spool, to direct alternating fluidflow to the slave valve spool as a result of the linear motion of thepilot valve spool to thereby induce linear motion of the slave valvespool; and a fixed slave valve sleeve surrounding and concentric withthe slave valve spool to receive alternating fluid flow from the movableslave valve spool to direct the output flow of hydraulic fluid to thefluid flow control output ports of the servo valve, the slave valvespool and the slave valve sleeve providing a power amplification stagefor the servo valve output flow that minimizes the volume of trappedhydraulic fluid passing from the pilot valve spool and the pilot valvesleeve to the output ports of the servo valve; the servo valve-actuatedtest assembly further comprising a vibration test fixture having a sliptable and a piston actuator affixed to the slip table, the output flowof hydraulic fluid from the control output ports of the servo valvedirected to the piston for applying a reciprocating linear motion to thepiston; and in which the output flow from the control outlet ports ofthe servo valve has an output frequency and the reciprocating motion ofthe piston is transferred to the slip table to induce vibrational motionin the slip table at a frequency corresponding to the output frequencyof the output flow of fluid from the servo valve.
 2. Apparatus accordingto claim 1 in which the energy input to the servo valve comprises avoice coil for a applying reciprocating linear motion to the pilot valvespool at controlled frequencies, including means for producingalternating output flow from the control output ports of the servo valvein response to the linear reciprocating motion of the pilot spool. 3.Apparatus according to claim in which the slip table is vibrated atfrequencies in the range from about 1000 Hz. to about 2000 Hz.
 4. A highfrequency linear hydraulic servo valve-actuated vibrating test assemblyfor producing alternating hydraulic fluid flow control outputs the servovalve-actuated test assembly comprising:an energy input sourcecomprising a transducer to receive an alternating input signal of aselected frequency and convert the input signal to linear mechanicalvibrational motion; a servo valve having a pair of fluid flow controloutput ports and in which an output flow of hydraulic fluid is producedfrom the control output ports of the servo valve; the servo valveincluding a movable hydraulic pilot valve spool that receives thevibrational motion from the transducer to apply linear reciprocatingmotion to the pilot valve spool along an axis; a movable hydraulic slavevalve spool surrounding and concentric with the pilot valve spool; afixed pilot valve sleeve surrounding and concentric with the pilot valvespool, the pilot valve sleeve located between the pilot valve spool andthe slave valve spool, to direct alternating fluid flow to the slavevalve spool as a result of the linear motion of the pilot valve spool tothereby induce linear motion of the slave valve spool; and a fixed slavevalve sleeve surrounding and concentric with the slave valve spool toreceive alternating fluid flow from the movable slave valve spool todirect the output flow of hydraulic fluid to the fluid flow controloutput ports of the servo valve, the slave valve spool and the slavevalve sleeve providing a power amplification stage for the servo valveoutput flow that minimizes the volume of trapped hydraulic fluid passingfrom the pilot valve spool and the pilot valve sleeve to the outputports of the servo valve; the servo valve-actuated test assembly furthercomprising a vibration test fixture having a fixture base for carrying aunit under test, and a hydraulic vibration actuator affixed to thefixture base and comprising a pair of opposed cylinders, each cylinderhaving a separate piston that reciprocates along a linear axis in acorresponding bore within the cylinder, each piston having acorresponding piston rod, the piston rods extending in oppositedirections from the actuator, each piston having an end face oppositefrom the piston's corresponding piston rod, said end face exposed totrapped hydraulic fluid contained within a volume in the bore adjacentthe end face of each piston, each cylinder having a separate inlet portto the trapped volume of fluid adjacent the piston, and in which thehydraulic servo valve output flow from the control output ports of theservo valve are connected to the inlet ports of the pistons forsupplying hydraulic fluid to the trapped volumes within the pistoncylinders for reciprocating the piston rods to induce a linear vibratingmotion to the fixture base.
 5. Apparatus according to claim 4 which thepistons and cylinders are spaced apart and the servo valve is located ina space between the pistons with the control output ports of the servovalve in close proximity to the inlet ports of the cylinders so as tominimize the volume of hydraulic fluid flow from the servo valve to thepiston cylinders.
 6. Apparatus according to claim 5 in which thevibration test fixture base includes a horizontal slip table, and inwhich the spaced-apart pistons and the servo valve are affixed to abottom center portion of the slip table.
 7. Apparatus according to claim6 in which the end face of each piston is in pressure contact against asupport affixed to the fixture base, with matching spherically curvedbearing surfaces between each piston rod end and corresponding support,to allow for relative sliding and rotating motion between the bearingsurfaces.
 8. Apparatus according to claim 4 which the energy input tothe servo valve comprises a voice color applying reciprocating linearmotion to the pilot spool at controlled frequencies, including means forproducing alternating output flow from the control output ports of theservo valve in response to the linear reciprocating motion of the pilotspool.
 9. Apparatus according to claim 6 in which the slip table isvibrated at frequencies in the range from about 1000 Hz. to about 2000Hz.